Fluid lubricated bearing device

ABSTRACT

An inner space of a housing sealed with a seal member as well as internal pores of a bearing member (pores in a porous structure) are filled with a lubricating oil without the presence of air, so that the oil surface of the lubricating oil is within a seal space. Under a reduced pressure of 100 Torr, no lubricating oil leaks outside of the housing even at any attitude of the fluid lubricated bearing device such as normal, inverted, or horizontal attitude.

CROSS-REFERENCE TO RELATED APPLICATION

This application claims the priority benefit of Japan applicationsserial no. 2001-347725, filed Nov. 13, 2001, 2002-29520, filed Feb. 6,2002, 2002-35790, filed Feb. 13, 2002 and 2002-281599, filed Sep. 26,2002.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a fluid lubricated bearing device forallowing a lubricating oil film to be produced in a radial bearing gapto provide non-contact support of a rotating member. For example, thisbearing device is suitable for use with: spindle motors incorporatedinto information apparatus, for example, a magnetic disk device such asan HDD or FDD, an optical disk device such as a CD-ROM, CD-R/RW, orDVD-ROM, a magneto-optical disk device such as an MD or MO; scannermotors incorporated into a copier, laser beam printer (LBP), or barcodereader; and small motors incorporated into electric appliances such asan axial fan.

2. Description of the Related Art

Each of the motors of the aforementioned types is required to satisfyhigher speed, lower cost, and lower acoustic noise requirements inaddition to higher rotational accuracy. One of the components that aredecisively responsible for these requirements is a bearing forsupporting the spindle of these motors. In recent years, use of thosefluid lubricated bearings have been attempted or actually made whichhave sufficiently good properties to satisfy the aforementionedrequirements. The fluid lubricated bearings of these types are largelydivided into two categories: one with means for generating dynamicpressure to provide dynamic pressure to the lubricating oil in bearinggaps or the so-called hydrodynamic bearing, and the other without meansfor generating dynamic pressure or the so-called fluid cylindricalbearing (with its bearing surface shaped in a perfect circle).

FIG. 7 is an example configuration of a spindle motor of an informationapparatus incorporating a hydrodynamic bearing device 31. This spindlemotor, used for a disk drive device of a DVD-ROM or the like, isprovided with a fluid lubricated bearing device 31 for rotatablysupporting an axial member 32, a support member 34 attached to the axialmember 32 to support an object to be driven such as an optical disk 33(a turntable in this illustrated example), and a motor stator 35 andmotor rotor 36 which are opposed to each other via a radial gap presenttherebetween.

The fluid lubricated bearing device 31 mainly consists of a housing 21having an opening portion at one end and a bottom portion at the otherend, a cylindrical bearing member 22 secured to the inner peripheralsurface of the bearing member 22, the axial member 32 inserted into theinner peripheral surface of the bearing member 22, a thrust plate 23providing on the bottom portion of the housing 21, and a seal member 24attached to the opening portion of the housing 21. On the innerperipheral surface of the bearing member 22 or the outer peripheralsurface of the axial member 32, there are provided grooves forgenerating dynamic pressure (dynamic pressure generating grooves). Onthe other hand, a lubricating oil is loaded into the space inside thehousing 21.

The stator 35 is attached to the outer periphery of the housing 21 ofthe fluid lubricated bearing device 31, while the rotor 36 is attachedto the support member 34. Flowing an electric current through the stator35 causes the rotor 36 to rotate due to exciting force establishedbetween the stator 35 and the rotor 36, thereby allowing the supportmember 34 and the axial member 32 to rotate integrally.

The rotation of the axial member 32 causes the dynamic pressuregenerating grooves to produce a dynamic pressure action of thelubricating oil in the radial bearing gap between the inner peripheralsurface of the bearing member 22 and the outer peripheral surface of theaxial member 32, thereby providing radial non-contact support to theouter peripheral surface of the axial member 32. In addition, the endsurface of the other end of the axial member 32 (the lower side in FIG.7) is supported in the direction of thrust by means of the thrust plate23.

In addition to the one mentioned above, such an bearing device is alsoavailable for providing non-contact support to the axial member in theradial and thrust direction by means of the dynamic pressure actionproduced in both the radial and thrust bearing gaps. In general, in thebearing of this type, a space defined by a thrust carrying face formedon the bottom portion of the housing and an end face of the axial memberopposed thereto is formed in a sealed arrangement. Accordingly, on theouter periphery of the axial member, there are formed axial grooves(circulation grooves) provided with openings at both the end facesthereof to expose this seal space to ambient air.

The lubricating oil is loaded into the space inside of the housing 21generally at the time of assembly of a spindle motor before the axialmember 32 is installed, and the axial member 32 is built therein afterthe lubricating oil has been loaded. For this reason, it cannot beavoided to ingest ambient air into the inner space of the housing 21.Additionally, the air inside the inner space of the housing may bethermally expanded or contracted due to variations in ambienttemperature, heat generated by the motor, or variations in atmosphericpressure during operations at high altitudes or transportation byairfreight. This may cause the lubricating oil to be pushed out of theseal space defined by the inner peripheral surface of the seal member 24and the outer peripheral surface of the axial member 32 to migrateoutside thereof. In particular, when the motor is operated in itsinverted position (i.e., the opening portion of the housing 21 isoriented downwardly) or in its horizontal position (i.e., the openingportion of the housing 21 is oriented horizontally), the lubricating oilmay readily flow toward the opening portion and stay there, therebyincreasing the likelihood of leakage of the lubricating oil.

For the reasons mentioned above, motors incorporating the conventionalfluid lubricated bearing device 31 were unreliable during operations intheir inverted and horizontally oriented positions, and thus limitationwas imposed on their operable positions.

Furthermore, in the fluid lubricated bearing device 31 configured asdescribed above, the thrust bearing portion employs the end face at theother end of the axial member 32 by means of the thrust plate 23, andthe axial member 32 is pushed against the thrust plate 23 by means ofthe magnetic force established between the stator 35 and the rotor 36,thereby restricting the axial movement of the axial member 32 towardsthe one end (the upper side in FIG. 7). However, when a shock loadexceeding the aforementioned magnetic force is imposed on the motor orthe motor is operated in an inverted or horizontally oriented position,the axial member 32 may move axially towards the one end relative to thehousing 21 to dislodge from the housing 21.

In the manufacturing process of the bearing member, a sleeve-shapedsintered metal is sized in a die to form the bearing member in apredetermined size. After the sizing followed by the removal from thedie, the outer periphery of the bearing member expands radiallyoutwardly due to spring-back. However, the circulation groove portion isnot in contact with the die during the sizing and therefore has not beenpressurized radially inwardly, thereby being provided with a less amountof spring-back when compared with other portions after the bearingmember has been removed from the die. Accordingly, as shown in FIG. 8,after the sizing, the outer and inner periphery of the bearing member 22is not of a complete round but of an uneven cross section with thevicinity of the circulation grooves 25 being reduced in diameter.Conventionally, the circulation grooves 25 are frequently formed at twopositions on the outer periphery of the bearing member 22 (at oppositepositions different by 180 degrees). In this case, the resulting crosssection obtained after the sizing displays an ellipse with its minoraxis being in the direction connecting between the circulation grooves25.

However, such an elliptical shape, if remains unchanged, causes a narrowportion (in the direction of the minor axis) and a wide portion (in thedirection of the major axis) to be formed in the radial bearing gapbetween the inner periphery of the bearing member 22 and the outerperiphery of the axial member 32. In this case, the lifting effectprovided by the hydrodynamic pressure on the shaft is reduced in thewide portion of the radial bearing gap, thereby causing axial runout dueto a degradation in bearing rigidity in the direction of the ellipticalmajor axis and thus probably having an adverse effect on NRRO or thelike.

An object of the present invention is to provide a fluid lubricatedbearing device and a motor that incorporates the same, which can beoperated or transported with stability at any attitude withoutlubricating oil leakage out of the housing due to expansion andcontraction of air remaining in the inner space thereof, under high andlow temperature environments, or reduced pressure environments duringoperations at high altitudes or transportation by airfreight.

Another object of the present invention is to restrict the axialmovement of the axial member relative to the housing towards one endthereof, thereby preventing the axial member from dislodging from thehousing.

Another object of the present invention is to eliminate differences inbearing rigidity in all directions that is caused by deformations of thebearing member after having undergone spring-back, thereby ensuring highrotational accuracy.

SUMMARY OF THE INVENTION

To achieve the aforementioned objects, the present invention provides afluid lubricated bearing device including: a housing with an openingportion provided on one end and a bottom portion provided on the otherend; an axial member and a bearing member accommodated in the housing; aradial bearing portion, provided between an inner peripheral surface ofthe bearing member and an outer peripheral surface of the axial member,for allowing an oil film of a lubricating oil produced in a radialbearing gap to provide radial non-contact support to the axial member;and a seal member arranged in the opening portion of the housing. Thefluid lubricated bearing device is configured such that the inner spaceof the housing is filled with the lubricating oil to such a level as notto cause leakage of the lubricating oil outside thereof even in thepresent of expansion or contraction of air remaining in the inner spaceof the housing under an environment of a reduced pressure. In theforegoing, barometric pressures under the environment of the reducedpressure are, for example, the atmospheric pressure to a pressure of 100Torr.

For example, the fluid lubricated bearing device configured as describedabove can be obtained by vacuuming the inner space of the housing andthen exposing the inner space of the housing to the atmospheric pressureto thereby replacing it with the lubricating oil (by vacuum pressureimpregnation). More specifically, the fluid lubricated bearing device isfirst assembled without a lubricating oil having been loaded therein(e.g., in the manners shown in FIGS. 1 to 4). Then, the entire or partof the fluid lubricated bearing device (at least the opening portionthat communicates with outside of the fluid lubricated bearing device)is soaked in the lubricating oil within a vacuum chamber, under whichthe air in the inner space of the housing is exhausted to create avacuum therein. Thereafter, the fluid lubricated bearing device isexposed to the atmospheric pressure to allow the lubricating oil to fillin the inner space of the housing. It is thus possible to provide thefluid lubricated bearing device according to the present invention.

However, depending on the degree of vacuum in the vacuum chamber, thereremains a small amount of air inside the housing after having beenexposed to the atmospheric pressure. With a large amount of airremaining in the housing, expansion or contraction of the remaining airdue to variations in ambient temperature would cause the lubricating oilto be pushed out of the housing, possibly resulting in leakage of thelubricating oil. In particular, the motor operating at an inverted orhorizontal attitude would cause the lubricating oil to flow in the innerspace of the housing to stay on the side of the opening portion, therebyresulting in the aforementioned lubricating oil leakage. Even with asmall amount of air remaining in the housing, the remaining air may beexpanded under environments of reduced pressures, such as operations athigh altitudes or transportation by airfreight, to push the lubricatingoil out of the housing, thereby possibly resulting in leakage of thelubricating oil.

The factors causing air to thermally expand include temperature andbarometric pressure. However, a calculation of the amount of airexpanded or contracted within the range of the temperature andbarometric pressure conceivable for operational environments tells thatthe pressure has greater effects on the expansion of air.

In general, the following environments under which a small spindle motorincorporating the fluid lubricated bearing device according to thepresent invention is operated and stored are often employed.

Temperature: Operating temperature 0 to 60° C., Storage temperature −40to 90° C.Barometric pressure: Atmospheric pressure to 0.3 atm duringtransportation (at an altitude of about 10,000 m)The rate of expansion is calculated in accordance with the equation ofstate of gas as follows:

PV=nRT

where P is the pressure, V is the volume, n is the number of moles, R isthe gas constant, and T is the absolute temperature.(1) For changes in temperature from −40 to 90° C. at a constantpressure:

V ₉₀ /V ₋₄₀=363/233=1.56 times

(2) For changes in pressure from the atmospheric pressure to 0.3 atm ata constant temperature:

V ₉₀ /V ₄₀=1/0.3=3.33 times

As can be seen from the foregoing, to suppress lubricating oil leakagedue to expansion of air, it is desirable to design the fluid lubricatedbearing device in consideration of changes in barometric pressure,having a greater effect than temperature, under the environmentsspecified within the range mentioned above in order to configure thedevice so as to be free from lubricating oil leakage.

For example, assume that the device is transported by airfreight at analtitude of 10,000 m. Since the barometric pressure is about 230 Torr(0.3 atm) in this case, it is necessary to load the lubricating oil soas not to cause lubricating oil leakage under an environment of areduced pressure of 230 Torr. It is desirable to check for lubricatingoil leakage at a pressure of 100 Torr with a certain allowance at thetime of inspection during fabrication of the bearing device.

As described above, the fluid lubricated bearing device according to thepresent invention and a motor incorporating the fluid lubricated bearingdevice can be operated or transported with stability irrespective of theattitude of the motor without lubricating oil leakage out of the housingdue to expansion and contraction of air remaining in the inner spacethereof, under high and low temperature environments, or reducedpressure environments during operations at high altitudes ortransportation by airfreight.

The fluid lubricated bearing device in which the inner space of thehousing is filled with the lubricating oil as described above is justlike an injector configured to be plugged at its top. This prevents theaxial member from being axially displaced due to vibrations duringtransportation or from being dislodged from the housing, to some extent.

Furthermore, to implement the aforementioned objects, the presentinvention provides a fluid lubricated bearing device including: ahousing with an opening portion provided on one end and a bottom portionprovided on the other end; an axial member and a bearing memberaccommodated in the housing; a radial bearing portion, provided betweenan inner peripheral surface of the bearing member and an outerperipheral surface of the axial member, for allowing an oil film of thelubricating oil produced in a radial bearing gap to provide radialnon-contact support to the axial member; a thrust bearing portion,provided on the bottom portion of the housing, for supporting an endface of the axial member on the same side as the other end of thehousing in a direction of thrust; and a seal member arranged in theopening portion of the housing. The fluid lubricated bearing device isconfigured such that a projected portion is provided on the axialmember, in contact with the seal member, to prevent an axial movement ofthe axial member relative to the housing towards the one end thereof.

In this configuration, the “projected portion” can be formed by beingintegrated with the axial member or by securing a member separate fromthe axial member to the axial member. On the other hand, the “projectedportion” is not limited in shape to a particular one, and may be formedin any shape such as annular, partially annular, point-like, or pin-likeshape. When the axial member is subjected to an external force or thegravitational force and thereby moved axially relative to the housingtowards one end thereof, the projected portion contacts the seal memberto thereby restrict further axial movement of the axial member relativeto the housing. This allows the axial member to be retained within thehousing and prevents it from dislodging therefrom.

In addition to the aforementioned configuration, it is also possible toprovide a configuration which allows the inner space of the housing tobe filled with the lubricating oil and which causes no lubricating oilleakage out of the housing even in the presence of expansion orcontraction of air remaining in the inner space of the housing underenvironments of reduced pressures from the atmospheric pressure to 100Torr.

In the aforementioned configuration, it is possible to provide an axialgap of 0.05 mm to 0.5 mm between the projected portion and the sealmember. The values of the axial gap are those for the end face of theother end of the axial member in contact with the thrust bearingportion.

In steady operation (or rotation with the end face of the other end ofthe axial member being supported in contact with the thrust bearingportion), it is necessary to provide a certain axial gap therebetween inorder to prevent the projected portion from contacting the seal member.In consideration of dimensional tolerances and assembly errors in eachcomponent, this axial gap needs to be 0.05 mm or more.

On the other hand, when the bearing device is subjected to repetitivevibrations or impact loads during operation or transportation, the axialmember can move axially relative to the housing within the range of theaforementioned axial gap due to the presence of the aforementioned axialgap. For this reason, when the aforementioned axial gap is excessivelylarge, the axial movement of the axial member relative to the housingcan cause outside air to be ingested into the housing through a sealspace (defined between the inner peripheral surface of the seal memberand the outer peripheral surface of the axial member) or the lubricatingoil inside the housing to be pushed out of the aforementioned sealspace. Furthermore, as the aforementioned axial gap becomes greater, alarger amount of the lubricating oil is filled in the inner space of thehousing, thereby providing larger variations in volume of thelubricating oil due to thermal expansion or contraction thereof. It istherefore necessary to provide a larger volume to the aforementionedseal space so as to accommodate the variations in volume to therebyprevent lubricating oil leakage out of the housing. However, it is oftendifficult to increase the axial size of the seal member due to aspace-wise restriction. Additionally, it is not preferable to increasethe inner diameter of the seal member since this can lead to a reductionin sealing function (degradation in capillary force).

According to the results of tests described later, it has been observedthat the aforementioned axial gap less than or equal to 0.5 mm canprevent lubricating oil leakage from inside to outside of the housing.An appropriate range of the aforementioned axial gap is from 0.05 mm to0.5 mm, preferably from 0.05 mm to 0.3 mm.

In the aforementioned configuration, it is possible to provide a sealspace, tapered so as to gradually expand towards the one end, betweenthe inner peripheral surface of the seal member and the outer peripheralsurface of the axial member opposite thereto. The seal space, tapered asdescribed above, causes the lubricating oil in the seal space to bedrawn towards the narrower side of the seal space (into the housing) bycapillary force. This prevents the lubricating oil leakage from insideto outside of the housing.

The tapered seal space described above can be formed by providing atapered face on at least one of the inner peripheral surface of the sealmember and the outer peripheral surface of the axial member. In theconfiguration with the tapered face provided on the outer peripheralsurface of the axial member, the lubricating oil in the seal space issubjected to centrifugal force during rotation of the axial member,thereby causing the lubricating oil to be drawn towards the narrowerside of the seal space along the tapered face of the axial member(towards inside of the housing). Thus, since the lubricating oil isretained in place by the centrifugal force in addition to theaforementioned capillary force, it is further ensured to prevent leakageof the lubricating oil.

Now, consider the fluid lubricated bearing device configured so as tosupport the axial end portion of the axial member in the direction ofthrust in contact with the thrust bearing portion provided on the bottomportion of the housing. This configuration may allow the lubricating oilto increase in pressure in the space around the thrust bearing portion,thereby causing a difference in pressure to occur between it and thelubricating oil in the seal space between the inner peripheral surfaceof the seal member and the outer peripheral surface of the axial member.This difference in pressure can also likely occur when the dynamicpressure generating grooves are axially asymmetrically formed in widthon the radial bearing portion as well as when the dynamic pressuregenerating grooves are designed symmetrically but with a large machiningtolerance (in the taper shape on the axial member or the innerperipheral surface of the bearing member and dimensional accuracy of thewidth of the dynamic pressure generating grooves).

Such a difference in pressure would cause a local negative pressure tobe produced in the lubricating oil in the inner space of the housing tothereby generate air bubbles in the lubricating oil, resulting inlubricating oil leakage or vibrations. Additionally, an increase inpressure of the lubricating oil around the thrust bearing portion wouldcause the axial member to be lifted or a decrease in pressure on thethrust bearing portion side would cause the axial member to be pushedagainst the carrying member such as the thrust plate, thereby causingthe carrying member to be abnormally worn.

These problems can be eliminated by providing circulation grooves thatcommunicate the thrust bearing portion with the seal space. That is,even with the difference in pressure of the lubricating oil between thespace around the thrust bearing portion and the seal space, thelubricating oil can flow between both the spaces through the circulationgrooves, thereby equalizing the pressure of the lubricating oil betweenboth the spaces.

The aforementioned circulation grooves can be configured of a firstradial groove formed on the side of the bottom portion of the housingbetween one end face of the bearing member and a surface of the housingopposite thereto, a second radial groove formed on the side of theopening portion of the housing between the other end face of the bearingmember and a surface of the seal member opposite thereto, and an axialgroove formed between the outer peripheral surface of the bearing memberand the inner peripheral surface of the housing.

The “fluid lubricated bearing device” according to the present inventionincludes a so-called hydrodynamic bearing device with dynamic pressuregenerating means for generating a dynamic pressure in the lubricatingoil in the bearing gap, and a so-called fluid cylindrical bearing device(a bearing device having a bearing surface of a complete round in crosssection) without the dynamic pressure generating means. However, thehydrodynamic bearing device is preferable which has an advantage inshaft supporting function. For the hydrodynamic bearing device, as theaforementioned “dynamic pressure generating means”, it is possible toprovide a hydrodynamic bearing device configured to have dynamicpressure generating grooves on one of the inner peripheral surface ofthe bearing member and the outer peripheral surface of the axial member,which face to each other via the radial bearing gap. It is also possibleto provide a hydrodynamic bearing device configured to have one of theaforementioned circumferential surfaces formed in a non-circular shapeor a plurality of circular segments such as two-arc segments, three-arcsegments, or four-arc segments (a bearing with its radial bearingsurface formed of a plurality of circular segments is also called an“arc bearing”). In the former bearing device, as the shape of thedynamic pressure generating grooves, it is possible to employ a varietyof well known dynamic pressure generating groove shapes such asherringbone patterns, spiral patterns, or patterns of a plurality ofaxial grooves (a bearing with a plurality of axial grooves provided onits radial bearing surface is also called a “stepped bearing”).Furthermore, a thrust dynamic pressure bearing portion may be configuredby forming dynamic pressure generating grooves in herringbone patternsor in a spiral fashion on one of the surfaces facing to each other viathe thrust bearing gap.

On the other hand, as the material of the bearing member, it is possibleto employ copper alloys, stainless steel, brass, aluminum alloys or thelike in addition to porous sintered metal.

Furthermore, to achieve the aforementioned objects, the presentinvention provides a fluid lubricated bearing device including an axialmember, a bearing member formed of a sintered metal containing oil andopposed to an outer periphery of the axial member via a radial bearinggap, and a housing with the bearing member secured to an inner peripherythereof. The fluid lubricated bearing device is configured such that theaxial member is rotated relative to the bearing member to generate ahydrodynamic pressure in the radial bearing gap to provide non-contactsupport to the axial member, and grooves are provided with openings attheir both end faces, through which a lubricating fluid flows, on theouter periphery of the bearing member. The fluid lubricated bearingdevice is provided with three or more, or preferably three, of theaforementioned grooves.

Three or more of the grooves provided as described above can increasethe bearing rigidity in each direction, thereby providing improvedrotational accuracy to the bearing.

The axial member is provided with a flange portion facing to one endface of the bearing member to generate a hydrodynamic pressure in thethrust bearing gap formed between the end face of the bearing member andthe end face of the flange portion, thereby making it also possible toprovide non-contact support to the axial member in the direction ofthrust. In this case, a larger amount of the lubricating fluid in thethrust bearing gap flows into the grooves due to the effects ofcentrifugal force. However, the three or more circulation grooves canpositively accommodate such a lubricating fluid.

When the dynamic pressure generating grooves on the bearing surface areformed asymmetrically to push the lubricating fluid into theaforementioned thrust bearing gap, a much larger amount of thelubricating fluid flows into the grooves. However, even in this case,the lubricating fluid can be accommodated with an allowance.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view illustrating a hydrodynamic bearingdevice according to a first embodiment of the present invention;

FIG. 2 is a cross-sectional view illustrating a hydrodynamic bearingdevice according to a second embodiment of the present invention;

FIG. 3 is a cross-sectional view illustrating a hydrodynamic bearingdevice according to a third embodiment of the present invention;

FIG. 4 is a cross-sectional view illustrating a hydrodynamic bearingdevice according to a fourth embodiment of the present invention;

FIG. 5 is a cross-sectional view illustrating a hydrodynamic bearingdevice used for test;

FIG. 6 is view illustrating a heat cycle pattern;

FIG. 7 is a cross-sectional view illustrating a spindle motor thatincorporates a conventional hydrodynamic bearing device;

FIG. 8 is a cross-sectional view illustrating a two-arc segmentedbearing;

FIG. 9 is a cross-sectional view illustrating an example of dynamicpressure generating means in which a plurality of axial dynamic pressuregenerating grooves are formed on the inner peripheral surface of abearing member;

FIG. 10 is a cross-sectional view illustrating an example of dynamicpressure generating means in which the inner peripheral surface of abearing member is formed of a plurality of arc segments;

FIG. 11 is a cross-sectional view illustrating an example of dynamicpressure generating means in which the inner peripheral surface of abearing member is formed of a plurality of arc segments;

FIG. 12 is a cross-sectional view illustrating an example of dynamicpressure generating means in which the inner peripheral surface of abearing member is formed of a plurality of arc segments;

FIG. 13 is a cross-sectional view illustrating an example of a radialbearing portion formed in the shape of a complete round without dynamicpressure generating means;

FIG. 14 is a cross-sectional view illustrating a fluid lubricatedbearing device with circulation grooves according to an embodiment;

FIG. 15 is a cross-sectional view illustrating a fluid lubricatedbearing device with circulation grooves according to another embodiment;

FIG. 16 is a view showing the results of tests conducted under reducedpressures;

FIG. 17 is a view showing the results of tests conducted using an actualfluid lubricated bearing device;

FIG. 18 is a view showing the results of tests conducted on the leakageof the lubricating oil;

FIG. 19 is a view showing the results of tests conducted on the leakageof the lubricating oil;

FIG. 20 is a longitudinal sectional view illustrating a hydrodynamicbearing device according to the present invention;

FIG. 21 is a perspective view illustrating a bearing member;

FIG. 22 is a view illustrating the non-dimensional rigidity of variousbearings;

FIG. 23 is a cross-sectional view illustrating a three-arc segmentedbearing;

FIG. 24 is a cross-sectional view illustrating a four-arc segmentedbearing;

FIG. 25 is an explanatory view illustrating eccentricity; and

FIG. 26 is a longitudinal sectional view illustrating a dynamic bearingdevice in which an asymmetrical radial bearing surface is formed in theaxial direction.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Now, the present invention will be explained below in accordance withthe embodiments.

FIG. 1 is a cross-sectional view illustrating a hydrodynamic bearingdevice 1 as an example of a fluid lubricated bearing device according toa first embodiment of the present invention. The fluid lubricatedbearing device 1, incorporated into a spindle motor used, for example,in an information apparatus as shown in FIG. 7, consists mainly of acylindrical bottomed housing 2 with an opening portion 2 a provided onone end (at the upper side in FIG. 1) and a bottom portion 2 c providedon the other end (at the lower side in FIG. 1), a cylindrical bearingmember 3 secured to the inner peripheral surface of the housing 2, anaxial member 4, and a seal member 5 secured to the opening portion 2 aof the housing 2. As described later, a first radial bearing portion R1and a second dynamic pressure bearing portion R2 are provided in anaxially spaced-apart relation between the inner peripheral surface 3 aof the bearing member 3 and the outer peripheral surface 4 a of theaxial member 4. On the other hand, there is also provided a thrustbearing portion T between the bottom portion 2 c of the housing 2 andthe lower end face 4 b of the axial member 4.

The housing 2 is formed, for example, of a soft metal such as brass, andis provided with a cylindrical side portion 2 b and the bottom portion 2c. For example, there is provided a plastic thrust plate 6 on a regionserving as a thrust bearing surface at the inner bottom surface of thebottom portion 2 c. In this embodiment, the housing 2 is configured suchthat the side portion 2 b is integrated with the bottom portion 2 c.Alternatively, the side portion 2 b and the bottom portion 2 c may beconfigured separately from each other such that a metallic lid-shapedmember serving as the bottom portion 2 c is press-fit into the openingportion on the other end of the side portion 2 b and then tightlysecured thereto with an adhesive or the like. In this case, the thrustplate 6 is disposed upon the aforementioned lid-shaped member.

The axial member 4 is formed, for example, of a stainless steel(SUS420J2) or the like, with the lower end face 4 b being formed in aspherically convex shape.

For example, the bearing member 3 is formed of a porous sintered metal,especially of a porous sintered metal mainly composed of copper. On theother hand, on the inner peripheral surface 3 a of the bearing member 3,there are provided two regions or upper and lower regions, which serveas radial bearing surfaces (the radial bearing surfaces of the firstradial bearing portion R1 and the second radial bearing portion R2) inan axially spaced-apart relation. For example, in these regions, thereare provided dynamic pressure generating grooves 3 a 1 and 3 a 2 havingherringbone patterns, respectively.

The axial member 4 is inserted into the inside of the bearing member 3along the inner peripheral surface 3 a thereof. As a result, the outerperipheral surface 4 a of the axial member 4 faces to the respectiveregions (two or the upper and lower regions) serving as the radialbearing surfaces on the inner peripheral surface 3 a of the bearingmember 3 via the radial bearing gaps. On the other hand, the lower endface 4 b of the axial member 4 contacts the upper face of the thrustplate 6.

The seal member 5, annular in shape, is press-fit into the innerperipheral surface of the opening portion 2 a of the housing 2, and thensecured thereto with an adhesive or the like. In this embodiment, theinner peripheral surface 5 a of the seal member 5 is formedcylindrically with the lower end face 5 b of the seal member 5 being incontact with the upper end face 3 b of the bearing member 3.

The inner peripheral surface 5 a of the seal member 5 faces to the outerperipheral surface 4 a of the axial member 4 via a predetermined gap,thereby forming a cylindrical seal space S1 therebetween. The innerspace of the housing 2 sealed with the seal member 5 as well as theinternal pores of the bearing member 3 (pores in the porous structure)are filled with a lubricating oil without the presence of air, so thatthe oil surface of the lubricating oil is within the seal space S1. Theseal space S2 is set so as to have a volume greater than the amount ofvariations in volume, caused by changes in temperature within itsallowable operational range, of the lubricating oil filled in the innerspace of the housing 2. This allows the oil surface of the lubricatingoil to be always maintained within the seal space S2 even in thepresence of a variation in volume of the lubricating oil caused by achange in temperature.

For example, the lubricating oil is loaded into the inner space of thehousing 2 in the following manner. First, each of the parts (the housing2, the bearing member 3, the axial member 4, the thrust plate 6, and theseal member 5) is assembled into the fluid lubricated bearing device 1having no lubricating oil loaded therein, which is in turn soaked in alubricating oil within a vacuum chamber. This allows the air in theinner space of the housing 2 to be drawn and evacuated due to the vacuuminside the vacuum chamber, thereby leaving no air in the inner space.Thereafter, when the vacuum chamber is exposed to the atmosphericpressure, the inner space of the housing 2 is filled with thelubricating oil. After the lubricating oil has been introduced therein,the fluid lubricated bearing device 1 is taken out of the vacuumchamber, and then heated to the operational upper-limit temperature ofthe fluid lubricated bearing device 1. As the fluid lubricated bearingdevice 1 is heated, the lubricating oil filled in the inner space of thehousing 2 is thermally expanded to discharge excessive lubricating oilout of the seal space S1. This allows the oil surface of the lubricatingoil to be maintained within the seal space S1 even when the fluidlubricated bearing device 1 operates at the operational upper-limittemperature. After the heating is stopped, the oil surface of thelubricating oil is lowered as the temperature decreases to settle to anappropriate level within the seal space S1.

In the aforementioned lubricating oil loading process, a slight amountof air may be left in the inner space of the housing 2 depending on thedegree of vacuum in the vacuum chamber. However, such an amount ofremaining air may be acceptable that is controlled to a predeterminedlevel sufficiently enough not to allow the lubricating oil to be pushedout of the seal space S1 and thereby migrate outside of the housing 2due to the expansion of the air remaining in the inner space of thehousing 2 under the potential environments of operation andtransportation of the fluid lubricated bearing device 1 and the motorincorporating the same. In this embodiment, under a reduced pressure of100 Torr, lubricating oil leakage to outside of the housing 2 iseliminated when the fluid lubricated bearing device 1 is placed in anormal position (with the side of the opening portion 2 a of the housing2 being oriented upwardly), in an inverted position (with the side ofthe opening portion 2 a of the housing 2 being oriented downwardly), ina horizontal position (with the side of the opening portion 2 a of thehousing 2 being oriented horizontally), and in a tilted position (withthe side of the opening portion 2 a of the housing 2 being oriented in atilted direction).

In the fluid lubricated bearing device 1 configured as described above,rotation of the axial member 4 causes a dynamic pressure to be producedin the lubricating oil in the aforementioned radial bearing gap, therebyallowing the film of the lubricating oil formed in the aforementionedradial bearing gap to provide radially rotatable non-contact support tothe outer peripheral surface 4 a of the axial member 4. This allows thefirst radial bearing portion R1 and the second radial bearing portion R2to be formed that provide radially rotatable non-contact support to theaxial member 4. At the same time, the thrust plate 6 provides contactsupport to the lower end face 4 b of the axial member 4, thereby formingthe thrust bearing portion T for rotatably supporting the axial member 4in the direction of thrust.

The fluid lubricated bearing device 1 according to this embodimenteliminates lubricating oil leakage from inside of the housing 2 tooutside thereof and enables stable operation and transportationirrespective of the position of the motor even in the presence ofexpansion or contraction of air remaining inside of the housing due tovariations in ambient temperature, heat generated in the motor, orvariations in atmospheric pressure during operations at high altitudesor transportation by airfreight.

FIG. 2 illustrates a hydrodynamic bearing device 1′ according to asecond embodiment. The fluid lubricated bearing device 1′ according tothis embodiment is different from the aforementioned first embodiment inthat a seal space S2 defined between the inner peripheral surface of aseal member 5′ and the outer peripheral surface of an axial member 4′facing to the seal member 5′ is tapered so as to gradually expand(outwardly) towards one end of the housing 2. In this embodiment, toform the tapered seal space S2, the inner peripheral surface of the sealmember 5′ is provided with a tapered face 5 a′ that is increasedgradually in diameter towards one end, while the outer peripheralsurface of the axial member 4′ facing to the tapered face 5 a′ isprovided with a tapered face 4 a 1′ that is reduced gradually indiameter towards the one end. One of the tapered face 5 a′ and thetapered face 4 a 1′ can be provided with a cylindrical surface.

As shown in the enlarged view circled with alternate long and shortdashed lines in FIG. 2, the oil surface of the lubricating oil L presentwithin the seal space S2 causes the lubricating oil L within the sealspace S2 to be drawn due to capillary force towards the narrower side ofthe seal space S2 (towards the other end or inside of the housing 2).This effectively prevents leakage of the lubricating oil L from insideto outside of the housing 2. Additionally, the tapered face 4 a 1′provided on the outer peripheral surface of the axial member 4′ causesthe lubricating oil L in the seal space S2 to be subjected tocentrifugal force during rotation of the axial member 4′ and therebydrawn along the tapered face 4 a 1′ towards the narrower side of theseal space S2 (towards inside of the housing 2). Thus, since thelubricating oil L is retained in place by the centrifugal force inaddition to the aforementioned capillary force, it is further ensured toprevent leakage of the lubricating oil L when compared with the fluidlubricated bearing device 1 according to the aforementioned firstembodiment.

FIG. 3 illustrates a hydrodynamic bearing device 1 according to a thirdembodiment. The fluid lubricated bearing device 1, incorporated into aspindle motor used, for example, in an information apparatus as shown inFIG. 7, consists mainly of a cylindrical bottomed housing 2 with anopening portion 2 a provided on one end (at the upper side in FIG. 3)and a bottom portion 2 c provided on the other end (at the lower side inFIG. 3), a cylindrical bearing member 3 secured to the inner peripheralsurface of the housing 2, an axial member 4, and a seal member 5 securedto the opening portion 2 a of the housing 2. As described later, a firstradial bearing portion R1 and a second dynamic pressure bearing portionR2 are provided in an axially spaced-apart relation between the innerperipheral surface 3 a of the bearing member 3 and the outer peripheralsurface 4 a of the axial member 4. On the other hand, there is alsoprovided a thrust bearing portion T between the bottom portion 2 c ofthe housing 2 and the lower end face 4 b of the axial member 4.

The housing 2 is formed, for example, of a soft metal such as brass, andis provided with a cylindrical side portion 2 b and the bottom portion 2c. For example, there is provided a plastic thrust plate 6 on a regionserving as a thrust bearing surface at the inner bottom surface of thebottom portion 2 c. In this embodiment, the housing 2 is configured suchthat the side portion 2 b is integrated with the bottom portion 2 c.Alternatively, the side portion 2 b and the bottom portion 2 c may beconfigured separately from each other such that a metallic lid-shapedmember serving as the bottom portion 2 c is press-fit into the openingportion on the other end of the side portion 2 b and then tightlysecured thereto with an adhesive or the like. In this case, the thrustplate 6 is disposed upon the aforementioned lid-shaped member.

The axial member 4 is formed, for example, of a stainless steel(SUS420J2) or the like, with the lower end face 4 b being formed in aspherically convex shape. A disk-shaped washer 7 or a projected portionis press-fit over the outer peripheral surface 4 a of the axial member 4and secured thereto with appropriate means such as an adhesive.

For example, the bearing member 3 is formed of a porous sintered metal,especially of a porous sintered metal mainly composed of copper. On theother hand, on the inner peripheral surface 3 a of the bearing member 3,there are provided two regions or upper and lower regions, which serveas radial bearing surfaces (the radial bearing surfaces of the firstradial bearing portion R1 and the second radial bearing portion R2) inan axially spaced-apart relation. For example, in these regions, thereare provided dynamic pressure generating grooves 3 a 1 and 3 a 2 havingherringbone patterns, respectively.

The axial member 4 is inserted into the inside of the bearing member 3along the inner peripheral surface 3 a thereof. As a result, the outerperipheral surface 4 a of the axial member 4 faces to the respectiveregions (two or the upper and lower regions) serving as the radialbearing surfaces on the inner peripheral surface 3 a of the bearingmember 3 via the radial bearing gaps. On the other hand, the lower endface 4 b of the axial member 4 contacts the upper face of the thrustplate 6.

The seal member 5, annular in shape, is press-fit into the innerperipheral surface of the opening portion 2 a of the housing 2, and thensecured thereto with an adhesive or the like. In this embodiment, theinner peripheral surface 5 a of the seal member 5 is formedcylindrically, and the lower end face 5 b of the seal member 5 isaxially spaced apart from the upper end face 3 b of the bearing member 3by a predetermined axial gap X to face thereto.

The washer 7 provided on the axial member 4 is disposed within the axialgap X. With the lower end face 4 b of the axial member 4 being incontact with the upper face of the thrust plate 6, the upper end face 7a of the washer 7 is spaced apart from the lower end face 5 b of theseal member 5 by an axial gap X1 while the lower end face 7 b of thewasher 7 is spaced apart from the upper end face 3 b of the bearingmember 3 by an axial gap X2. The axial gap X1 is 0.05 mm to 0.5 mm inlength, preferably 0.05 mm to 0.3 mm. The axial gap X2 may be set at alength that is sufficiently enough not to allow the lower end face 7 bof the washer 7 to contact the upper end face 3 b of the bearing member3 during rotation of the axial member 4, preferably at 0.05 mm or morein consideration of dimensional tolerances and assembly errors in eachcomponent. The axial gap X2 may be equal to the axial gap X1 in length,or greater or less than the axial gap X1.

The inner peripheral surface 5 a of the seal member 5 faces to the outerperipheral surface 4 a of the axial member 4 via a predetermined gap,thereby forming a cylindrical seal space S1 therebetween. The innerspace of the housing 2 sealed with the seal member 5 as well as theinternal pores of the bearing member 3 (pores in the porous structure)are filled with the lubricating oil without the presence of air, so thatthe oil surface of the lubricating oil is within the seal space S1. Theseal space S1 is set so as to have a volume greater than the amount ofvariations in volume, caused by changes in temperature within itsallowable operational range, of the lubricating oil filled in the innerspace of the housing 2. This allows the oil surface of the lubricatingoil to be always maintained within the seal space S1 even in thepresence of a variation in volume of the lubricating oil caused by achange in temperature.

For example, as in the same manner as the first embodiment, thelubricating oil is loaded into the inner space of the housing 2. Thus,even when the air remaining in the inner space of the housing isexpanded or contracted under the environments between the atmosphericpressure and a reduced pressure of 100 Torr, lubricating oil leakagefrom inside of the housing 2 is eliminated irrespective of the attitudeof the motor.

In this embodiment, when the axial member 4 is subjected to an externalforce or gravitational force to move axially towards the one endrelative to the housing 2, this causes the washer 7 provided on theaxial member 4 to contact the seal member 5, thereby restricting furtheraxial movement of the axial member 4. This allows the axial member 4 tobe always retained within the housing 2 and thereby prevented fromdislodging from the housing 2.

Furthermore, since the washer 7 is separated from the seal member 5 byan axial gap X1 of 0.05 mm to 0.5 mm, the washer 7 and the seal member 5are not in contact with each other under normal operations (duringrotation with the lower end face 4 b of the axial member 4 beingsupported in contact with the thrust plate 6), thereby providing stablerunning conditions. Additionally, even when the axial member 4 movesaxially within the range of the axial gap X1 relative to the housing 2,air will never be ingested into the housing 2 or the lubricating oilfilled in the housing 2 will never be pushed out of the seal space S1 toleak therefrom.

The other discussions follow those of the first embodiment and will notbe repeatedly presented here.

FIG. 4 illustrates a fluid lubricated bearing device 1′ according to afourth embodiment. The fluid lubricated bearing device 1′ according tothis embodiment is different from the aforementioned third embodiment inthat a seal space S2 defined between the inner peripheral surface of aseal member 5′ and the outer peripheral surface of an axial member 4′facing to the seal member 5′ is tapered so as to gradually expand(outwardly) towards one end of the housing 2. In this embodiment, toform the tapered seal space S2, the inner peripheral surface of the sealmember 5′ is provided with a tapered face 5 a′ that is increasedgradually in diameter towards one end, while the outer peripheralsurface of the axial member 4′ facing to the tapered face 5 a′ isprovided with a tapered face 4 a 1′ that is reduced gradually indiameter towards the one end. One of the tapered face 5 a′ and thetapered face 4 a 1′ can be provided with a cylindrical surface.

As shown in the enlarged view circled with alternate long and shortdashed lines in FIG. 4, the oil surface of the lubricating oil L presentwithin the seal space S2 causes the lubricating oil L within the sealspace S2 to be drawn due to capillary force towards the narrower side ofthe seal space S2 (towards the other end or inside of the housing 2).This effectively prevents leakage of the lubricating oil L from insideto outside of the housing 2. Additionally, the tapered face 4 a 1′provided on the outer peripheral surface of the axial member 4′ causesthe lubricating oil L in the seal space S2 to be subjected tocentrifugal force during rotation of the axial member 4′ and therebydrawn along the tapered face 4 a 1′ towards the narrower side of theseal space S2 (towards inside of the housing 2). Thus, since thelubricating oil L is retained in place by the centrifugal force inaddition to the aforementioned capillary force, it is further ensured toprevent leakage of the lubricating oil L when compared with the fluidlubricated bearing device 1 according to the aforementioned thirdembodiment.

In the embodiments described above, the dynamic pressure generatinggrooves 3 a 1 and 3 a 2 having herringbone patterns are formed asdynamic pressure generating means on the inner peripheral surface 3 a ofthe bearing member 3 serving as the radial bearing surfaces (the radialbearing surfaces of the first radial bearing portion R1 and the secondradial bearing portion R2). However, instead of the herringbone pattern,dynamic pressure generating grooves having spiral patterns may beformed. Alternatively, as shown in FIG. 9, a plurality of dynamicpressure generating grooves 3 a 3 formed axially may be provided asdynamic pressure generating means on the inner peripheral surface 3 a ofthe bearing member 3 serving as the radial bearing surface (a so-called“stepped bearing”).

Alternatively, as shown in FIGS. 10 to 12, as dynamic pressuregenerating means, the inner peripheral surface 3 a of the bearing member3 serving as the radial bearing surfaces (the radial bearing surfaces ofthe first radial bearing portion R1 and the second radial bearingportion R2) may be formed non-circularly, for example, of a plurality ofarc-shaped segments (a so-called “arc-shaped bearing”). The exampleshown in FIG. 10 is configured such that the inner peripheral surface 3a of the bearing member 3 is formed of two arc segmented faces (3 a 4and 3 a 5). The center of curvature O1 of the arc segmented face 3 a 4and the center of curvature O2 of the arc segmented face 3 a 5 areoffset at an equal distance from the outer peripheral surface 4 a of theaxial member 4 (having a complete round shape), respectively. Theexample shown in FIG. 11 is configured such that the inner peripheralsurface 3 a of the bearing member 3 is formed of three arc-shapedsegments (3 a 6, 3 a 7, and 3 a 8). The center of curvature O3 of thearc segmented face 3 a 6, the center of curvature O4 of the arcsegmented face 3 a 7, and the center of curvature O5 of the arcsegmented face 3 a 8 are offset at an equal distance from the outerperipheral surface 4 a of the axial member 4 (having a complete roundshape), respectively. The example shown in FIG. 12 is configured suchthat the inner peripheral surface 3 a of the bearing member 3 is formedof four arc-shaped segments (3 a 9, 3 a 10, 3 a 11, and 3 a 12). Thecenter of curvature O6 of the arc segmented face 3 a 9, the center ofcurvature O7 of the arc segmented face 3 a 10, the center of curvatureO8 of the arc segmented face 3 a 11, and the center of curvature O9 ofthe arc segmented face 3 a 12 are offset at an equal distance from theouter peripheral surface 4 a of the axial member 4 (having a perfectcircular shape), respectively.

The aforementioned dynamic pressure generating means may be provided onthe outer peripheral surface 4 a of the axial member 4.

Alternatively, as shown in FIG. 13, the first radial bearing portion R1(the second radial bearing portion R2) may be a “cylindrical bearing”having no dynamic pressure generating means.

The embodiment shown in FIG. 14 is the hydrodynamic bearing deviceaccording to the embodiments, shown in FIGS. 1 and 2, in which thethrust bearing portion T is in communication with the seal space S1,defined between the inner peripheral surface 5 a of the seal member 5and the outer peripheral surface 4 a of the axial member 4, throughcirculation grooves 10 that are provided at one or a plurality ofpositions (at two positions in the illustrated example) in thecircumferential direction.

The inner peripheral surface 5 a of the seal member 5 faces to the outerperipheral surface 4 a of the axial member 4 via a predetermined gap,thereby forming a cylindrical seal space S1 therebetween. The innerspace of the housing 2 sealed with the seal member 5 as well as theinternal pores of the bearing member 3 (pores in the porous structure)are filled with the lubricating oil without the presence of air, so thatthe oil surface of the lubricating oil is within the seal space S1. Theseal space S1 is set so as to have a volume greater than the amount ofvariations in volume, caused by changes in temperature within itsallowable operational range, of the lubricating oil filled in the innerspace of the housing 2. This allows the oil surface of the lubricatingoil to be always maintained within the seal space S1 even in thepresence of a variation in volume of the lubricating oil caused by achange in temperature.

The circulation grooves 10 has first and second radial grooves 10 a and10 c and an axial groove 10 b, with both the radial grooves 10 a and 10c being coupled to both ends of the axial groove 10 b. The first radialgroove 10 a is formed between one end of the bearing member 3 or an endface 3 c (closer to the bottom portion 2 c of the housing) and a face ofthe housing 2 opposite thereto, more specifically, an inner face 2 c 1of the bottom portion 2 c. On the other hand, the second radial groove10 c is formed between the other end of the bearing member 3 or the endface 3 b (closer to the opening portion 2 a of the housing) and a faceof the seal member 5 opposite thereto, more specifically, the inner face5 b of the seal member 5. The axial groove 10 b is formed between theouter peripheral surface of the bearing member 3 and the innerperipheral surface of the side portion 2 b of the housing 2.

In the embodiment shown in FIG. 14, the first and second radial grooves10 a and 10 c are formed on the end faces 3 c and 3 b of the bearingmember 3, respectively, while the axial groove 10 b is formed on theouter peripheral surface of the bearing member 3. When the axial member4 is rotated to cause the lubricating oil to increase in pressure in thespace of the thrust bearing portion T (in the space around the axial endportion of the axial member 4), this causes the lubricating oil to startflowing towards the seal space S1 from the vicinity of the thrustbearing portion T, thereby allowing the lubricating oil to be equalizedin pressure in the vicinities of the thrust bearing portion T and theseal space S1. This prevents generation of air bubbles caused by anegative pressure locally produced in the lubricating oil as well aslubricating fluid leakage or vibrations resulting therefrom.Furthermore, the axial member 4 is also prevented from being lifted dueto an increase in pressure of the lubricating oil in the vicinity of thethrust bearing portion T. When the pressure in the seal space S1 hasincreased on the contrary to the foregoing, likewise, the circulationgrooves 10 allows the lubricating oil to be maintained at an equalpressure in the vicinities of the thrust bearing portion T and the sealspace S1, thereby making it possible to avoid lubricating oil leakagecaused by air bubbles generated or a harmful effect such as abnormalabrasion of the thrust plate 6 which may be caused by the axial member 4being pushed against the bottom portion 2 c of the housing.

FIG. 15 illustrates an embodiment in which circulation grooves 10′ areformed on members opposite to the bearing member 3 (the housing 2 andthe seal member 5). That is, the first radial groove 10 a′ is formed onthe inner face 2 c 1 of the bottom portion 2 c of the housing, thesecond radial groove 10 c′ is formed on the lower end face 5 b′ of theseal member 5, and the axial groove 10 b′ is formed on the innerperipheral surface of the side portion 2 b of the housing. Thecirculation grooves 10′ can provide the same effects as the embodimentshown in FIG. 14.

FIG. 14 illustrates the cylindrical seal space S1 and FIG. 15illustrates the tapered seal space S2. However, a seal space is notlimited to one having a particular shape. Thus, on the contrary totheses examples, the tapered seal space S2 can be employed in theembodiment of FIG. 14 while the cylindrical seal space S1 can beemployed in the embodiment of FIG. 15.

EXAMPLE

Bearing devices of five types to be tested (examples 1 to 2 andcomparative examples 1 to 3) were fabricated by loading a lubricatingoil in the aforementioned manner (by vacuum pressure impregnation) intothe fluid lubricated bearing devices 1, configured as shown in FIG. 1,with the vacuum chamber being kept at different levels of vacuum tothereby have different amounts of air remaining in the inner space ofthe housing 2 after exposed to the atmospheric pressure. It is difficultto measure the amount of air remaining in the inner space of the housingafter the lubricating oil has been loaded therein by vacuum pressureimpregnation. However, for example, it can be estimated that air of 50vol % of the inner space volume remains inside the housing after exposedto the atmospheric pressure when the vacuum chamber is reduced inpressure to 380 Torr (half the atmospheric pressure). This approach wasused to estimate the amount of air remaining in the housing inner space.

Each of the aforementioned test bearing devices was tested to check forlubricating oil leakage when left under an environment of a reducedpressure (the reduced-pressure test) and when incorporated into anactual motor to operate being turned ON and OFF at various operatingpositions under the atmospheric pressure (the motor test). The resultsof the tests are shown in FIG. 16 (the reduced-pressure test) and FIG.17 (the motor test). The test conditions are as follows.

Reduced-Pressure Test

Degree of vacuum: 100 Torr

Motor Test

Motor used: CD-ROM motor for actual use

Rotational speed: 8000 rpm

Ambient temperature: 60° C.

Motor attitude: Normal, Horizontal, Inverted

Running condition: ON-OFF (30 seconds per cycle)

Test time: 300,000 cycles

According to the reduced-pressure test, since different amounts of airremained in the inner space of the housing were dependent on the degreeof vacuum of the vacuum chamber, some test bearing devices leakedlubricating oil under reduced pressures even though having had undergonevacuum pressure impregnation (comparative examples 1 to 3).

According to the motor test, those test bearing devices whoselubricating oil was loaded in drops (comparative examples 2 and 3)leaked lubricating oil during 50,000 to 200,000 cycles at the horizontaland inverted positions. On the other hand, those test bearing deviceshaving undergone vacuum pressure impregnation (examples 1 and 2, andcomparative example 1) leaked no lubricating oil up to 300,000 cycles atany position.

Thus, by loading the lubricating oil so as not to leak it even under areduced pressure of 100 Torr as in the examples, it is possible toprovide a fluid lubricated bearing device that can be operated andtransported in a stable state without lubricating oil leakage even inany potential operational attitude or any environmental condition.

On the other hand, with the configuration shown in FIG. 3, three typesof fluid lubricated bearing devices 1 were fabricated which had an axialgap X1, between the washer 7 and the seal member 5, of 0.1 mm, 0.3 mm,and 0.5 mm, respectively. A dummy disk 9 equivalent to a practical loadwas placed over the axial member 4 of each of the fluid lubricatedbearing devices 1 (see FIG. 5), which was in turn subjected to a dropimpact test at 1000 G and then checked for lubricating oil leakage frominside of the housing 2. The drop impact of 1000 G was determined withreference to the property of resistance to impact load which is requiredof a spindle motor employed for recent portable-use apparatus such as anHDD device for use in notebook computers. Furthermore, the test wasconducted on the conventional fluid lubricated bearing device shown inFIG. 7 under the same conditions as described above (comparative example4). The test results are shown in FIG. 18.

As can be seen from the results shown in FIG. 18, during the applicationof an impact load of 1000 G, the axial member dislodged from the housingin the comparative example 4 (shaft dislodgement), whereas in theexamples 3 to 5, no shaft dislodgement occurred and no lubricating oilleakage was observed.

Furthermore, the fluid lubricated bearing devices according to theaforementioned examples 3 to 5 and comparative example 4 wereincorporated into an actual motor (a polygon scanner motor for use witha laser beam printer), respectively, and then operated under thefollowing conditions to check for lubricating oil leakage from inside ofthe housing. The test results are shown in FIG. 19.

Running Conditions

Actual motor: Polygon scanner motor for LBP

Rotational speed: 30,000 rpm

Heat cycle pattern: See FIG. 6

Test time: 20 cycles

Motor attitude: Horizontal and inverted

As can be seen from the test results shown in FIG. 19, during operationin heat cycles, lubricating oil leakage was observed in the comparativeexample 4, whereas no lubricating oil leakage was observed at any of thehorizontal and inverted attitudes in the examples 3 to 5.

As is obvious from the foregoing, the present invention provides thefollowing effects.

(1) The inner space of the housing is filled with the lubricating oil tosuch a level as not to cause leakage of the lubricating oil outsidethereof even in the present of expansion and contraction of airremaining in the inner space of the housing under environments ofreduced pressures, particularly, from the atmospheric pressure to apressure of 100 Torr. This prevents leakage of the lubricating oil frominside to outside of the housing and thereby allows stable operation andtransportation of the fluid lubricated bearing device at any attitudesuch as normal, inverted, or horizontal attitude under any environmentalconditions, conceivable as the environments for operation andtransportation of the motor incorporating the fluid lubricated bearingdevice, such as high and low temperature environments or reducedpressure environments such as during operations at high altitudes ortransportation by airfreight.

(2) The inner space of the housing is filled with the lubricating oilwithout air involved therein, thereby making it possible to preventlubricating oil leakage or cavitation caused by ingestion of air.

(3) The axial member is provided with a projected portion in contactwith the seal member to restrict the axial movement of the axial memberrelative to the housing towards one end thereof, thereby allowing theaxial member to be always retained within the housing and prevented frombeing dislodged from the housing.

(4) An axial gap of 0.05 mm to 0.5 mm is defined between the projectedportion and the seal member, thereby making it possible to providestable operation without the projected portion in contact with the sealmember. At the same time, it is also possible to prevent ingestion ofair into the housing or lubricating oil leakage from inside of thehousing even when the axial member moves axially within the range of theaxial gap relative to the housing.

(5) A seal space, tapered so as to gradually expand towards one end, isprovided between the inner peripheral surface of the seal member and theouter peripheral surface of the axial member opposite thereto, therebyproviding improved sealing function to prevent lubricating oil leakagemore effectively.

(6) Circulation grooves are provided to communicate the thrust bearingportion with the seal space, thereby equalizing the pressures in thethrust bearing portion and the seal space even in the presence of adifference in pressure of the lubricating oil between the thrust bearingportion and the seal space. This makes it possible to preventdetrimental effects such as generation of air bubbles, lubricating oilleakage, lifting of the shaft, and abnormal wear of the thrust plate.

FIG. 20 illustrates a hydrodynamic bearing device according to anotherembodiment. The hydrodynamic bearing device according to this embodimentmainly consists of a sleeve-shaped bearing member 3, an axial member 4,and a bottomed housing 2 cylindrical in shape.

The bearing member 3 is formed of sintered meal impregnated with alubricating oil or a lubricating grease to fill fine pores inside thesintered metal with oil.

For example, the sintered metal is mainly composed of copper, iron, orboth, preferably with 20 to 95% of copper. As in the prior art, thebearing member 3 is fabricated through the processes of powder pressuremolding, sintering, sizing, and impregnation of oil. The bearing member3 obtained through those processes is provided by means of stamping orthe like, on its inner peripheral surface and one end face 3 c, withgrooves 3 a 1 and 3 a 2 for generating dynamic pressure (dynamicpressure generating grooves), described later.

On the inner periphery of the bearing member 3, there are formed radialbearing surfaces 12 a and 12 b that have a plurality of dynamic pressuregenerating grooves 3 a 1 and 3 a 2 as dynamic pressure generating means.In the illustrated example, two radial bearing surfaces 12 a and 12 bare axially spaced apart from each other. However, the number of theradial bearing surfaces 12 a, 12 b is not limited to two but may be one,or three or more. The dynamic pressure generating grooves 3 a 1 and 3 a2 of the radial bearing surfaces 12 a and 12 b are good enough if theyare tilted at angles relative to the shaft, and may be arranged inherringbone patterns as illustrated or in a spiral fashion. On the otherhand, it is also possible to employ a non-complete-round radial bearingsurface, which has no dynamic pressure generating grooves, having aharmonic waveform or the like.

The axial member 4, formed of metal such as stainless steel, consists ofa straight axial portion 4 c and a disk-shaped flange portion 4 dprovided at an end of the axial portion 4 c. The axial portion 4 c andthe flange portion 4 d may be formed separately by press fitting orintegrally by means of forging or the like.

The housing 2, bottomed and formed cylindrically, has an opening at oneend with the other end closed. The bearing member 3 is secured to theinner periphery of the housing 2 by press fitting or with an adhesive orthe like. At this time, the axial portion 4 c of the axial member 4 isplaced on the inner periphery of the bearing member 3 while the flangeportion 4 d is placed in the space between the bottom portion 2 c of thehousing 2 and the one end face 3 c of the bearing member 3. Asillustrated, the housing bottom portion 2 c is integrated with thecylindrical side portion 2 b or may be formed separately from the sideportion 2 b and then fit therein. The opening portion of the sideportion 2 b is sealed with the seal member 5 to prevent leakage of oilserving as a lubricating fluid. There is provided a slight axial gapbetween the seal member 5 and the end face 3 b of the bearing member 3opposite thereto in order to provide improved oil-retaining effects.

Under this condition, both end faces 4 d 1 and 4 d 2 of the flangeportion 4 d face to the one end face 3 c of the bearing member 3 and athrust carrying surface 9 a 1 of the housing bottom portion 2 c,respectively. On the end face 3 c of the bearing member 3 and the innerbottom face 2 c 1 of the housing bottom portion 2 c, which face to theflange portion 4 d, there are formed thrust bearing surfaces 11 a and 11b that have a plurality of dynamic pressure generating grooves (notshown) as the dynamic pressure generating means, respectively. Thedynamic pressure generating grooves of the thrust bearing surfaces 11 aand 11 b may be arbitrarily formed in shape. It is possible to formdynamic pressure generating grooves in the herringbone patterns as inthe radial bearing surfaces 12 a and 12 b or in a spiral fashion as wellas a stepped thrust bearing surface. The dynamic pressure generatinggrooves may be formed on both the end faces 4 d 1 and 4 d 2 of theflange portion 4 d instead of the axial member end face 3 c or the innerbottom face 2 c 1. In these cases, the aforementioned thrust bearingsurface is formed on both the end faces 4 d 1 and 4 d 2 of the flangeportion 4 d.

There is an oil serving as a lubricating fluid filled in a fine gap(radial bearing gap) between the radial bearing surfaces 12 a and 12 band the outer peripheral surface of the axial portion 4 c and in a finegap (thrust bearing gap) between the thrust bearing surfaces 11 a and 11b and surfaces opposite thereto (the end faces 4 d 1 and 4 d 2 of theflange portion 4 d in the illustrated example), respectively. Duringrelative rotation between the axial member 4 and the bearing member 3(rotation of the axial member 4 relative thereto in this embodiment),the operation of each of the radial bearing surfaces 12 a, 12 b, 11 a,and 11 b causes a dynamic pressure in the oil in the radial bearing gapand the thrust bearing gap, thereby providing non-contact support to theaxial member 4 in both the radial and thrust directions relative to thebearing member 3.

On the outer periphery of the bearing member 3, there are formed axialgrooves having openings at their both end faces 3 b and 3 c orcirculation grooves 10. The circulation grooves 10 communicate a sealedspace defined between the bottom portion 2 c of the housing 2 and theend face 3 c of the bearing member 3 with outside of the bearing,serving as a passage for allowing the oil to flow axially. The oil inthe circulation grooves 10 is absorbed by the bearing member 3, andexudes from the surface of the bearing member 3 to be supplied again toeach bearing gap. Because of the reasons described later, the presentinvention provides circumferentially three or more, preferably three,circulation grooves 10 at equal intervals (see FIG. 21).

When three circulation grooves 10 are formed in such a manner, as shownin FIG. 23, the bearing member 3 is deformed into a shape generallytriangular in cross section with three circular segments 13 of anincreased radius (a bearing member deformed in this manner ishereinafter referred to as a three-arc segmented bearing) after sizingdue to the difference in the amount of spring-back between thecirculation grooves 10 and other portions. On the other hand, when fourcirculation grooves 10 are formed, as shown in FIG. 24, the bearingmember is deformed into a shape generally rectangular in cross sectionwith four circular segments 13 of an increased radius (hereinafterreferred to as a four-arc segmented bearing) after sizing for the samereason. Although not illustrated, in the cases where five or morecirculation grooves 10 are provided, the bearing member is deformed intoa shape generally polygonal in cross section with circular segments ofan increased radius being the same in number as the circulation grooves10 (a five-arc segmented bearing, six-arc segmented bearing and soforth). For simplicity, FIGS. 23 and 24 exaggerate the degree ofdeformation relative to a complete round; however, the deformations arenot as distinct as illustrated to the human eyes.

FIG. 22 shows the results of non-dimensional rigidity analysis of an oilfilm from a two-arc segmented bearing, a three-arc segmented bearing,and a four-arc segmented bearing. These results were obtained by using acomputer to numerically solve a second-order differential equation orthe Reynolds equation by which the pressure of the fluid in the bearinggap was expressed. In the regions of negative pressures, Reynoldsconditions were used as pressure boundary conditions. As used herein,the Reynolds conditions mean those that provide a pressure gradient ofzero at discontinued portions of oil film and satisfy the continuity offlow.

The two-arc segmented bearing, the three-arc segmented bearing, and thefour-arc segmented bearing described above are provided with two, three,and four circulation grooves 10 of a width of 10 degrees at equalintervals in the circumferential direction, respectively. Additionally,the ratio of the axial length L of the bearing member 3 to the outerdiameter D (L/D) is set at 0.5 for all the bearings. Furthermore, theeccentricity E of the axial member 4 is based on E=0.1 (however,E=0.0868 for the two-arc segmented bearing). As shown by a dashed linein FIG. 25, the centers of the bearing member 3 and the axial member 4coincide with each other at E=0, while the axial member 4 is inscribedin the bearing member 3 at ε=1, as shown by a chain double-dashed linein FIG. 25 (in which the width of the radial bearing gap isexaggerated).

Kxx, Kyy, Kxy, and Kyx represent the elastic constants of the oil film,each obtained by integrating a numerically determined distribution ofpressure over the bearing surface and then numerically integrating theresulting x and y loads in the x- and y-directions, respectively. Theseconstants are expressed in no dimensions, and dimensional rigidity kijis expressed by the following equation in which the four non-dimensionalrigidities are expressed by Kij:

kij=(W/Cp)Kij

where W is the bearing load and Cp is the radial bearing gap.

A subscript “xx” represents X displacement causing a force in the Xdirection (in the direction of the minor axis of the ellipse), asubscript “yy” represents Y displacement causing a force in the Ydirection (in the direction of the major axis of the ellipse), asubscript “xy” represents Y displacement causing a force in the Xdirection, and a subscript “yx” represents X displacement causing aforce in the Y direction, respectively. Those with subscripts “xy” and“yx” are coupled items that show a force generated due to a displacementcaused by movement of not itself but others, and with those items beinggreater, the axial member 4 provides more unstable runout. As can beseen from FIG. 22, the two-arc segmented bearing keeps no balancebetween the Kxx and Kyy and therefore has a big difference in bearingrigidity depending on the load direction, whereas the three-arcsegmented bearing and the four-arc segmented bearing keep a balancetherebetween and therefore has no such a disadvantage. From theforegoing, the number of the circulation grooves 10 should be preferablythree or more so that the bearing member 3 approximates to the three-arcsegmented bearing or the four-arc segmented bearing in cross sectionafter spring-back.

On the other hand, consider how to control the radial bearing gap inmass production. The four-arc segmented bearing provides a bigdifference in inner diameter depending on the direction of measurement(see the arrows in FIG. 24), whereas the three-arc segmented bearingprovides a small difference (see the arrows in FIG. 23). For thisreason, when compared with the four-arc segmented bearing, the three-arcsegmented bearing can have a higher tolerance on the inner diameter andthus can be fabricated at lower cost. According to FIG. 22, thethree-arc segmented bearing is further preferable in that the three-arcsegmented bearing has a smaller absolute value of the coupled items Kxyand Kyx. On the other hand, bearings with five-arc segments or moretakes an approximately complete round in cross section after havingdeformed due to spring-back, thereby raising concerns that the axialmember 4 will be subjected to unstable self-excited vibrations calledwhirling as well as increasing machining costs of the grooves. For thereasons described above, it is the most preferable to form threecirculation grooves so that the bearing approximates to a three-arcsegmented bearing in cross section after having been deformed due tospring-back.

FIG. 26 illustrates a configuration in which the bearing surface 12 b,of the two radial bearing surfaces 12 a and 12 b, closer to the bottomportion of the housing 2, is formed axially asymmetrically relative tothe other bearing surface 12 a to enhance the force for driving the oiltowards the housing bottom portion through the dynamic pressuregenerating grooves 3 a 1 and 3 a 2. This may raise a concern that adisplacement of a region having oil film formed therein toward thehousing bottom portion causes an increased amount of oil to flow intothe circulation grooves 10, thereby making it impossible to sufficientlyaccommodate the flow rate of the oil through conventional twocirculation grooves (indicated by the reference numeral 25 in FIG. 8).However, three or more circulation grooves 10 can be provided asdescribed above, thereby avoiding such a drawback. The number of thecirculation grooves can be determined according to the flow of oil,however, it is the most preferable to form three or more circulationgrooves 10 as described above in consideration of rotational accuracy.

According to the present invention, since three or more circulationgrooves are formed on the outer periphery of the bearing member, it ispossible to prevent the bearing rigidity from becoming unstable becauseof the bearing member deformed due to spring-back after sizing andthereby provide improved rotational accuracy for the bearing.

1-2. (canceled)
 3. A fluid lubricated bearing device comprising: ahousing with an opening portion provided on one end and a bottom portionprovided on the other end; an axial member and a bearing memberaccommodated in the housing; a radial bearing portion, provided betweenan inner peripheral surface of the bearing member and an outerperipheral surface of the axial member, for allowing an oil film of alubricating oil produced in a radial bearing gap to provide radialnon-contact support to the axial member; a thrust bearing portion,provided on the bottom portion of the housing, for supporting an endface of the axial member on the same side as the other end of thehousing in a direction of thrust; and a seal member arranged in theopening portion of the housing, wherein a projected portion is providedon the axial member, in contact with the seal member, to prevent anaxial movement of the axial member relative to the housing towards theone end thereof.
 4. The fluid lubricated bearing device according toclaim 3, wherein an inner space of the housing is filled with thelubricating oil and no lubricating oil leakage out of the housing iscaused even in the presence of expansion or contraction of air remainingin an inner space of the housing under environments of reduced pressuresfrom the atmospheric pressure to 100 Torr.
 5. The fluid lubricatedbearing device according to claim 3, wherein an axial gap of 0.05 mm to0.5 mm is provided between the projected portion and the seal member. 6.The fluid lubricated bearing device according to claim 4, wherein anaxial gap of 0.05 mm to 0.5 mm is provided between the projected portionand the seal member. 7-11. (canceled)
 12. A motor comprising the fluidlubricated bearing device according to claims
 3. 13-14. (canceled)